1. Field of the Invention
The present invention relates generally to a valve timing control system for regulating or controlling valve open/close timing (also referred to simply as the valve timing) at which an intake valve and/or an exhaust valve of an internal combustion engine is opened and closed in dependence on operating state or condition of the engine. More particularly, the invention is concerned with a valve timing control system for controlling the actually detected valve timing so that it can converge rapidly or speedily to a desired valve by learning most retarded timing to be used as reference values for the control over a wide operation range of the engine while invalidating speedily information or data learned erroneously, to thereby prevent the quality of engine exhaust gas from deterioration.
2. Description of Related Art
In the technical field of the internal combustion engine, there has heretofore been well known a system for controlling variably the operation timing for at least one of an intake valve and an exhaust valve of the engine in dependence on the operation state thereof with a view to enhancing efficiency of intake and discharge operations of the engine cylinders.
In such control system, learned values (most retarded valve open/close timing indicative of angular position of the engine crank shaft serving as a reference for the control) in which dispersion among the products brought about in manufacturing the same has been canceled out are previously stored in a memory or a storage, for thereby allowing the valve open/close timing (control quantity) to be arithmetically determined in dependence on various operation states of the engine on the basis of the learned values. In other words, the learned values are always used in the valve open/close timing control. Thus, it is required that the learned values be determined to be optimal.
Typical ones of the hitherto known or conventional control system of the type mentioned above are disclosed, for instance, in Japanese Unexamined Patent Application Publications Nos. 299876/1994 (JP-A-6-299876) and 345264/1997 (JP-A-9-345264), respectively. For having better understanding of the concept underlying the present invention, description will first be made in some detail of a conventional valve timing control system for an internal combustion engine by reference to FIGS. 15 to 28.
FIG. 15 is a schematic diagram showing generally a configuration of a gasoline engine system equipped with a conventional valve open/close timing regulating mechanism. Referring to the figure, an internal combustion engine (hereinafter also referred to simply as the engine) denoted generally by reference numeral 1 constitutes a major part of the gasoline engine system and includes a plurality of cylinders (e.g. four cylinders). In FIG. 15, however, only one cylinder and associated components are illustrated representatively.
As is shown in FIG. 15, a cylinder block 2 forms a cylinder portion of the engine 1. A cylinder head 3 is connected to the cylinder block 2 at a top end thereof. A piston 4 is housed within each of cylinder chambers formed in the cylinder block 2 so as to move reciprocatively in the vertical direction. A crank shaft 5 is operatively coupled to the piston 4 at a bottom end thereof and caused to rotate in unison with the reciprocative motion of the piston 4.
A crank angle sensor 6 is disposed in the vicinity of the crank shaft 5. This sensor 6 may be constituted, for example, by an electromagnetic pickup or sensor designed for generating a crank angle signal SGT in synchronism with revolution of the engine 1. The crank angle signal SGT is utilized not only for detecting the engine speed or rotation number NE (rpm) of the engine 1 but also for detecting a predetermined reference crank angle (.degree. CA) of the crank shaft 5.
A signal rotor 7 is integrally mounted on the crank shaft 5 and has an outer peripheral surface formed with a pair of teeth 7a with an angular distance of 180.degree. therebetween. The teeth 7a are formed of a magnetic material. Thus, upon every passing of each tooth 7a in front of the crank angle sensor 6, the crank angle signal SGT of a pulse-like form is generated by the crank angle sensor 6.
A combustion chamber 8 is defined by the inner wall of the cylinder block 2 and those of the cylinder head 3 and a top wall of the piston 4. Air-fuel mixture charged into the engine 1 undergoes combustion within the combustion chamber 8. To this end, a spark plug 9 is disposed at the top of the cylinder head 3 so as to partially project into the combustion chamber 8. The air-fuel mixture is fired by electric discharge taking place at the spark plug 9.
A distributor 10 is installed, being operatively coupled to an exhaust-side cam shaft 20 (described hereinafter) which is mounted on the cylinder head 3. The distributor 10 is designed for applying a high firing voltage sequentially to the spark plugs 9 provided in the individual cylinders, respectively. An ignitor 11 is provided for generating the high ignition voltage.
More specifically, each of the spark plugs 9 is electrically connected to the distributor 10 by way of a high-voltage rated cord (not shown), wherein the high voltage outputted from the ignitor 11 is distributed to the individual spark plugs 9, respectively, by the distributor 10 in synchronism with the rotation of the crank shaft 5.
Further installed in association with the cylinder block 2 is a water temperature sensor 12 which serves for detecting the temperature W of cooling water flowing through a coolant passage. An intake port 13 is provided at an intake side of the cylinder head 3 while an exhaust port 14 is disposed at an exhaust side of the cylinder head 3. An intake passage 15 is communicated to the intake port 13 with an exhaust passage 16 being communicated to the exhaust port 14. An intake valve 17 is disposed in the intake port 13 of the cylinder head 3 while an exhaust valve 18 is installed in the exhaust port 14 formed in the cylinder head 3.
An intake-side cam shaft 19 is disposed above the intake valve 17 for driving the intake valve 17 to open and close the same. Similarly, the exhaust-side cam shaft 20 is disposed above the exhaust valve 18 for opening and closing the same.
An intake timing pulley 21 is mounted on the intake-side cam shaft 19 at one end thereof, while an exhaust timing pulley 22 is mounted on the exhaust-side cam shaft 20 at one end thereof. The intake timing pulley 21 and the exhaust timing pulley 22 are operatively coupled to the crank shaft 5 by means of a timing belt 23 so that each of the cam shafts 19 and 20 can rotate at a speed equal to a half of the rotation speed (rpm) of the crank shaft 5.
When the engine 1 is in the operating state, the driving torque of the crank shaft 5 is transmitted to the cam shafts 19 and 20, respectively, by way of the timing belt 23 and the timing pulleys 21 and 22 to thereby rotate the cam shafts 19 and 20, respectively.
Thus, the intake valve 17 and the exhaust valve 18 are driven, respectively, to the open/close states in synchronism with the rotation of the crank shaft 5 and hence the reciprocative motion of the piston 4. In other words, each of the intake valve 17 and the exhaust valve 18 is driven with a predetermined open/close timing in synchronism with a series of four strokes, i.e., the suction stroke, compression stroke, explosion (expansion) stroke and the exhaust stroke of the engine 1.
A cam angle sensor 24 is disposed in the vicinity of the intake-side cam shaft 19 and designed to generate a cam angle signal SGC for detecting the actuation timing (i.e., valve timing) of the intake valve 17.
A signal rotor 25 is integrally mounted on the intake-side cam shaft 19, wherein the outer peripheral surface of the signal rotor 25 is formed with four teeth 25a at an angular distance of 90.degree. therebetween. Each of the teeth 25a is formed of a magnetic material. Every time each of the teeth 25a passes in front of the cam angle sensor 24, a pulse-like cam angle signal SGC is generated by the cam angle sensor 24.
A throttle valve 26 is installed in the intake passage 15 at an intermediate portion thereof and adapted to be selectively opened or closed in response to actuation of an accelerator pedal (not shown), whereby the air-flow quantity (intake air flow) Q fed to the engine 1 is regulated. To this end, a throttle sensor 27 is operatively coupled to the throttle valve 26 for detecting the throttle opening degree .theta..
An intake air-flow sensor 28 is disposed at a location upstream of the throttle valve 26 for detecting the intake air flow Q in the intake passage 15 by resorting to, for example, thermal detection technique which per se is known in the art. Further, a surge tank 29 is installed at a location downstream of the throttle valve 26 for suppressing pulsation of the intake air flow Q.
A fuel injector 30 is mounted in the vicinity of the intake port 13 of each of the individual cylinders, respectively, for injecting fuel for charging the air-fuel mixture into the combustion chamber 8. Each of the fuel injectors 30 is ordinarily constituted by an electromagnetic valve which is opened upon electric energization thereof. Fuel is fed to each of the fuel injectors 30 under pressure from a fuel pump (not shown).
In operation of the engine 1, air is taken into the intake passage 15 while the fuel injector 30 injects the fuel in the direction toward the intake port 13. As a result of this, air-fuel mixture is produced in the intake port 13 to be sucked into the combustion chamber 8 through the intake valve 17 which is adapted to open in the suction stroke.
A variable valve timing mechanism (VVT) 40 is operatively coupled to the intake-side cam shaft 19 and designed to be driven hydraulically (by using lubricant oil of the engine 1) for changing or modifying the valve operation timing of the intake valve 17 (or at least one of the intake valve 17 and the exhaust valve 18). More specifically, the variable valve timing mechanism 40 is so designed as to regulate or change the valve operation timing of the intake valve 17 continuously by changing the angle of displacement of the intake-side cam shaft 19 relative to the intake timing pulley 21. To this end, an oil control valve (OCV) 80 is provided for supplying working oil to the variable valve timing mechanism 40 as well as for adjusting the amount of the working oil.
For the purpose of realizing overall control of the engine operation, there is provided an electronic control unit (hereinafter also referred to as the ECU) 100 for controlling operations of the various actuators such as the fuel injector 30, the ignitor 11, the oil control valve 80 and others for controlling the fuel injection quantity, the ignition timing, the valve operation timing and others on the basis of the output signals of the various sensors such as the signals indicating the intake air flow Q, the cooling water temperature W, the crank angle signal SGT, the cam angle signal SGC and others which represent the operation state of the engine. The electronic control unit 100 may be constituted by a microcomputer or microprocessor, as will be described later on.
Next, description is directed to a structure of a variable valve timing system including the variable valve timing mechanism 40 and the oil control valve 80 by reference to FIGS. 16 to 24. FIG. 16 is a side elevational view showing partially in section a structural arrangement around the intake-side cam shaft 19 provided in association with the variable valve timing mechanism 40. Further, this figure also shows a structure of the working oil supply means inclusive of the oil control valve 80 for driving the variable valve timing mechanism 40. Parenthetically, in FIG. 16, the components same as or equivalent to those described previously are denoted by like reference characters.
Referring to FIG. 16, the variable valve timing mechanism 40 serves to regulate or adjust the intake valve operation timing, while the oil control valve 80 controls the amount of working oil supplied to the variable valve timing mechanism 40. The intake timing pulley 21 rotates in synchronism with the crank shaft 5 through the medium of the timing belt 23 which moves in unison with the rotation of the crank shaft 5.
Transmitted translationally to the intake-side cam shaft 19 is the rotation of the intake timing pulley 21 with modified phase due to intervention of the variable valve timing mechanism 40.
A bearing 41 is fixedly mounted on the cylinder head 3 (see FIG. 15) for supporting rotatably the intake-side cam shaft 19.
A first oil passage 42 and a second oil passage 43 are provided in association with the intake-side cam shaft 19 and a rotor 52 (described hereinafter), respectively. The first oil passage 42 is communicated to a retarding hydraulic chamber 62 (also described hereinafter) for displacing angularly the rotor 52 in the retarding direction while the second oil passage 43 is communicated to an advancing hydraulic chamber 63 (also described hereinafter) for displacing angularly the rotor 52 in the advancing direction.
There is further provided an oil pump 91 for taking out the working oil (lubricating oil) from an oil pan 90 containing the working oil (lubricant). Additionally, an oil filter 92 is provided for purifying the working oil taken out from the oil pan 90. The oil pan 90, the oil pump 91 and the oil filter 92 cooperate to constitute the lubricating means for lubricating various parts or components of the engine 1 (see FIG. 15) and at the same time constitute a working oil supply means for the variable valve timing mechanism 40 in cooperation with the oil control valve 80.
A variety of sensors designated generally and collectively by reference numeral 99 includes the sensors such as the crank angle sensor 6 mentioned previously and others provided in association with the engine 1, wherein output signals of these sensors indicating various operation state information of the engine 1 are inputted to the electronic control unit 100.
A spool valve element 82 is mounted within the housing 81 of the oil control valve 80 to move slideably therein. A linear solenoid 83 controls the spool valve element 82 in accordance with a corresponding control signal supplied from the electronic control unit 100. A spring 84 is provided for urging resiliently the spool valve element 82 in the direction opposite to the driving direction of the linear solenoid 83.
The housing 81 is provided with various ports 85 to 87, 88a and 88b.
The oil supplying port 85 is hydraulically communicated to the oil pump 91 by way of the oil filter 92, wherein an A-port 86 is hydraulically communicated to the first oil passage 42 with a B-port 87 being communicated to the second oil passage 43. On the other hand, the exhaust ports 88a and 88b are hydraulically communicated to the oil pan 90.
When the engine 1 is operating, the working oil is discharged from the oil pan 90 by the oil pump 91 which is put into operation upon rotation of the crank shaft 5. The working oil as discharged is fed under pressure selectively to the first oil passage 42 or the second oil passage 43 by way of the oil filter 92 and the oil control valve 80.
The flow rate of oil in each of the first and second oil passages 42 and 43 is increased or decreased consecutively as the opening degrees of the ports 86 and 87 are changed due to sliding movement of the spool valve element 82. In this conjunction, it is noted that the opening degrees of the A-port 86 and the B-port 87 are determined, respectively, in dependence on the magnitude or value of the control current i (control quantity of the linear solenoid current) supplied to the linear solenoid 83. Hereinafter, this current will also be referred to as the linear solenoid current i.
The electronic control unit 100 controls the control current i on the basis of the signals outputted from the various sensors such as the crank angle sensor 6, the cam angle sensor 24 and others.
A housing 44 of the variable valve timing mechanism 40 is mounted rotatably relative to the intake-side cam shaft 19, wherein a casing 45 is fixedly secured to the housing 44. A back spring 46 which may be constituted by a leaf spring is disposed between a tip seal 49 (described hereinafter) and the casing 45 to resiliently urge the tip seal 49 against the rotor 52 (also described hereinafter).
A cover 47 is secured to the casing 45 by means of bolts 48 which secure the housing 44, the casing 45 and the cover 47 to one another. The tip seal 49 is pressed against the rotor 52 by means of the back spring 46 to thereby prevent flow of working oil between the hydraulic chambers partitioned from each other by the rotor 52 and the casing 45. A plate 50 is secured to the cover 47 by means of a screw 51.
The rotor 52 is fixedly mounted on the intake-side cam shaft 19 and disposed rotatably relatively to the casing 45. The rotor 52 is provided with a cylindrical holder 53 having a recess which is adapted to engage with a plunger 54 (described hereinafter).
The plunger 54 provided with a protrusion is adapted to move slideably within the housing 44 under the influence of resiliency of a spring 55 (described hereinafter) and a hydraulic pressure of the oil introduced into the holder 53. The spring 55 exerts a spring force for urging the plunger 54 toward the rotor 52. A plunger oil passage 56 feeds the working oil in the direction for applying the hydraulic pressure to the plunger 54 against the spring force of the spring 55. An air passage 57 is provided for setting constantly to the atmospheric pressure the space formed at the side of the plunger 54 at which the spring 55 is disposed.
The intake-side cam shaft 19 and the rotor 52 are connected fixedly to each other by means of a connecting bolt 58. On the other hand, the intake-side cam shaft 19 and the rotor 52 are interconnected by a rotatable shaft portion of a shaft member 59 which is mounted rotatably relative to the cover 47. An air passage 60 is formed so as to extend internally through the shaft member 59 and the intake-side cam shaft 19 for setting the inner space defined by the plate 50 to the atmospheric pressure.
FIG. 17 is a fragmentary sectional view showing a state in which a hydraulic pressure is applied to the plunger 54 by way of the plunger oil passage 56.
As can be seen in FIG. 17, the plunger 54 is forced to bear against the housing 44 with the spring 55 being compressed under the hydraulic pressure. As a result of this, the plunger 54 and the holder 53 are disengaged from each other to thereby allow the rotor 52 to rotate relative to the housing 44.
FIG. 18 is a sectional view taken along a line X--X in FIG. 16 as viewed in the direction indicated by arrows, FIG. 19 is a fragmentary sectional view for illustrating displacement of a slide plate 71, FIG. 20 is a sectional view taken along a line Y--Y in FIG. 16 as viewed in the direction indicated by arrows, and FIG. 21 is a sectional view taken along a line Z--Z in FIG. 16 as viewed in the direction indicated by arrows.
Referring to FIGS. 18 to 21, the bolts 48 are received screwwise in the bolt holes 61, respectively. There are provided sector-like retarding hydraulic chambers 62 for rotating first to fourth vanes 64 to 67, respectively, in a retarding direction, as described hereinafter. Parenthetically, these vanes 64 to 67 are formed integrally with the rotor 52.
Each of the retarding hydraulic chambers 62 is defined as enclosed by the rotor 52, the casing 45, the cover 47 and the housing 44 in correspondence to the first to fourth vanes 64 to 67, respectively. Further, the retarding hydraulic chambers 62 are hydraulically communicated with the first oil passage 42, being supplied with working oil from the first oil passage 42.
On the other hand, there are provided sector-like advancing hydraulic chambers 63 for rotating the first to fourth vanes 64 to 67 in an advancing direction. Each of the advancing hydraulic chamber 63 is defined, being enclosed-by the rotor 52, the casing 45, the cover 47 and the housing 44 in correspondence to the first to fourth vanes 64 to 67, respectively. Further, the advancing hydraulic chambers 63 are hydraulically communicated with the second oil passage 43, being supplied with the working oil from the second oil passage 43.
With the arrangement described above, the rotor 52 is displaced relative to the housing 44 in dependence on the amount of the working oil supplied to the retarding hydraulic chamber 62 and the advancing hydraulic chamber 63, whereby volumes of the retarding hydraulic chamber 62 and the advancing hydraulic chamber 63 are caused to change correspondingly.
The first vane 64 protrudes radially outwardly from the rotor 52. The holder 53 is fitted to the first vane 64 at the side facing the housing 44 with a communicating oil passage 70 (described hereinafter) being formed in the cover 47. A guide groove 72 (described hereinafter) is formed in an intermediate portion of each of the communicating oil passages 70. The plunger oil passage 56 extends through the holder 53 from a guide groove 72 to the housing 44.
Similarly, each of the second to fourth vanes 65 to 67 is so formed as to protrude from the rotor 52 outwardly in the radial direction. Further, a tip seal 73 (described hereinafter) is provided in a portion of each of the first to fourth vanes 64 to 67, which seal is brought into contact with the casing 45.
A vane supporting member 68 constitutes a center portion of the rotor 52. Shoes 69 are provided so as to protrude from the casing 45 inwardly in the radial direction. Each of the shoes 69 is provided with a bolt hole 61 for receiving the bolt 48 with the tip seal 49 being provided at a portion of the shoe 69 where the seal is brought into contact with the vane supporting member 68.
The communicating oil passage 70 is communicated with the retarding hydraulic chamber 62 and the advancing hydraulic chamber 63 formed at both sides of the first vane 64, respectively. The slide plate 71 is movable within the guide groove 72 (described hereinafter) formed in an intermediate portion of the communicating oil passage 70. The communicating oil passage 70 is divided or partitioned by the slide plate 71 so that no oil leakage can take place between the retarding hydraulic chamber 62 and the advancing hydraulic chamber 63.
With the arrangement described above, the slide plate 71 is caused to displace toward the advancing hydraulic chamber 63 when the hydraulic pressure within the retarding hydraulic chamber 62 is high (see FIG. 18). On the other hand, when the hydraulic pressure within the advancing hydraulic chamber 63 is high, the slide plate 71 is forced to move toward the retarding hydraulic chamber 62 (see FIG. 19).
As mentioned previously, the guide groove 72 is provided at an intermediate portion of the communicating oil passage 70, wherein an intermediate portion of the guide groove 72 is communicated to the plunger oil passage 56.
Thus, when the slide plate 71 moves toward the advancing hydraulic chamber 63 (see FIG. 18), the plunger oil passage 56 communicates with the retarding hydraulic chamber 62. Similarly, when the slide plate 71 moves toward the retarding hydraulic chamber 62 (see FIG. 19), the plunger oil passage 56 is set to the state communicating with the advancing hydraulic chamber 63.
The tip seal 73 is provided for each of the first to fourth vanes 64 to 67 for preventing occurrence of leakage of oil between the vanes and the casing 45. Parenthetically, arrows shown in FIGS. 18, 20 and 21 indicate the direction in which the variable valve timing mechanism 40 as a whole is rotated by means of the timing belt 23 and others.
Next, operations of the variable valve timing mechanism 40 and the oil control valve 80 will be described in the concrete.
At first, in the state in which the operation of the engine 1 is stopped, the rotor 52 assumes a most retarded position (i.e., the position at which the rotor 52 has been rotated to maximum relative to the housing 44 in the retarding direction), as is shown in FIG. 18.
In the state mentioned above, the hydraulic pressure of oil fed from the oil pump 91 to the oil control valve 80 is low (e.g. at the atmospheric pressure). Consequently, oil is supplied to neither the first oil passage 42 nor the second oil passage 43. Thus, no hydraulic pressure is applied to the plunger oil passage 56. As a result of this, the plunger 54 is resiliently pressed against the holder 53 under the force of the spring 55, as is shown in FIG. 16. In other words, the plunger 54 and the holder 53 engage with each other.
Upon starting of operation of the engine 1, the oil pump 91 is put into operation, whereby the hydraulic pressure supplied to the oil control valve 80 rises up. Consequently, hydraulic medium or oil is fed to the retarding hydraulic chamber 62 by way of the A-port 86. As a result of this, the slide plate 71 is caused to move toward the advancing hydraulic chamber 63 under the hydraulic pressure prevailing within the retarding hydraulic chamber 62, whereby hydraulic communication is established between the retarding hydraulic chamber 62 and the plunger oil passage 56. Thus, the plunger 54 is urged to displace toward the housing 44, which brings about disengagement between the plunger 54 and the rotor 52.
However, because the hydraulic pressure is applied to the advancing hydraulic chamber 63, each of the first to fourth vanes 64 to 67 remains in the state bearing on the shoes 69 in the retarding direction under hydraulic pressure. Accordingly, even when the plunger 54 is disengaged, the housing 44 and the rotor 52 are pressed against each other under the hydraulic pressure within the retarding hydraulic chambers 62, whereby vibration or shock can be suppressed to minimum.
Now, when the B-port 87 is opened for rotating the rotor 52 in the advancing direction, working oil is supplied to the advancing hydraulic chamber 63 by way of the second oil passage 43. Consequently, hydraulic pressure is transmitted to the communicating oil passage 70 from the advancing hydraulic chamber 63, as a result of which the slide plate 71 is caused to move toward the retarding hydraulic chamber 62 under the hydraulic pressure.
When the slide plate 71 is moved as mentioned above, the plunger oil passage 56 is placed in hydraulic communication with the advancing hydraulic chamber 63 by way of the communicating oil passage 70, whereby hydraulic pressure is transmitted to the plunger oil passage 56 from the advancing hydraulic chamber 63. Under the hydraulic pressure mentioned above, the plunger 54 is forced to move toward the housing 44 against the spring force exerted by the spring 55, which results in disengagement between the plunger 54 and the holder 53.
In this manner, by adjusting the amount of working oil by opening/closing the A-port 86 and the B-port 87 in the state where the plunger 54 and the holder 53 are disengaged from each other, rotation of the rotor 52 can be advanced or retarded relative to the rotation of the housing 44 owing to the regulation of the amount of working oil within the retarding hydraulic chamber 62 and the advancing hydraulic chamber 63.
Next, by referring to FIGS. 22 and 24, typical operation of the oil control valve 80 will be described. Incidentally, FIGS. 22 to 24 show, respectively, operating states of the oil control valve 80 when the control current i issued from the electronic control unit 100 assumes different values.
More specifically, FIG. 22 shows operation state of the oil control valve 80 when the value of the control current i is ia (e.g. 0.1 ampere) smaller than a reference current value ib (e.g. 0.5 ampere).
Referring to FIG. 22, the spool valve element 82 is resiliently urged to the left-hand side of the housing 81 under the force of the spring 84, whereby the oil supplying port 85 and the A-port 86 on one hand and the B-port 87 and the exhaust port 88b on the other hand are mutually communicated, as indicated by arrows.
In this state, working oil is fed to the retarding hydraulic chamber 62 while it is discharged from the advancing hydraulic chamber 63. Thus, the rotor 52 is forced to rotate in the counterclockwise direction relative to the housing 44 as indicated by arrow. Consequently, the phase of the intake-side cam shaft 19 retards relative to that of the intake timing pulley 21, whereby the intake valve 17 is placed in the retard control state, so to say.
FIG. 23 shows the operation state of the oil control valve 80 when the value of the control current i is equal to the reference current value ib (e.g. 0.5 ampere). In the state illustrated in FIG. 23, the forces exerted by the linear solenoid 83 and the spring 84, respectively, and acting in opposite directions to each other are in balance, as a result of which the spool valve element 82 is maintained at a position where both the A-port 86 and the B-port 87 are closed.
Consequently, the retarding hydraulic chamber 62 and the advancing hydraulic chamber 63 are in the states in which working oil is neither supplied nor discharged. Thus, the rotor 52 will be sustained at the current position so long as the leakage of working oil does not occur from the retarding hydraulic chamber 62 and the advancing hydraulic chamber 63, whereby the phase relation between the intake timing pulley 21 and the intake-side cam shaft 19 can be maintained in the currently prevailing state.
On the other hand, FIG. 24 shows operation state of the oil control valve 80 when the value of the control current i is ic (e.g. 1.0 ampere) which is greater than the reference current value ib (e.g. 0.5 ampere).
Referring to FIG. 24, the spool valve element 82 is urged to the right-hand side of the housing 81 under the influence of the linear solenoid 83, whereby the oil supplying port 85 and the B-port 87 on one hand and the A-port 86 and the exhaust port 88a on the other hand are mutually communicated, as indicated by arrows.
In this state, working oil is fed to the advancing hydraulic chamber 63 through the second oil passage 43, while being discharged from the retarding hydraulic chamber 62 through the first oil passage 42. Thus, the rotor 52 is forced to rotate in the clockwise direction relative to the housing 44 as indicated by arrow. Consequently, the phase of the intake-side cam shaft 19 is caused to advance relative to that of the intake timing pulley 21. Thus, the intake valve 17 is placed in the advance control state.
As can be seen from FIGS. 22 to 24, degree of hydraulic communication between the oil supplying port 85 and the A-port 86 or the B-port 87 as well as the degree of hydraulic communication between the exhaust port 88a or 88b and the A-port 86 or the B-port 87 can be controlled in dependence on the position of the spool valve element 82. In this conjunction, it is to be mentioned that the position of the spool valve element 82 and the value of the control current i flowing through the linear solenoid 83 bear a proportional relationship to each other.
FIG. 25 is a characteristic diagram for illustrating a relation between the value of the control current i flowing through the linear solenoid 83 and an actual valve timing change rate VTa. More specifically, there is graphically illustrated the rate of change of the actual valve timing (hereinafter also referred to as the actual valve timing change rate) designated by VTa as a function of the linear solenoid current i under predetermined operating condition of the engine 1. In FIG. 25, a positive or plus region of the actual valve timing change rate VTa represents the displacement in the advancing direction, while a negative or minus region of the actual valve timing change rate VTa corresponds to the region in which the displacement takes place in the retarding direction.
In FIG. 25, the electric current values ia, ib and ic represent values of the linear solenoid current i corresponding to the positions of the spool valve element 82 shown in FIGS. 22, 23 and 24, respectively. As can be seen from the figures, the value of the linear solenoid current i at which the actual valve timing Ta does not change (i.e., VTa=0) is only one current value ib at which the amount of working oil leaking from the hydraulic chambers 62 and 63 as well as the hydraulic pipe and the spool valve element 82 is in balance with the amount of working oil fed under pressure from the oil pump 91.
FIG. 26 is a characteristic diagram for illustrating variations in the relation between the control current i flowing through the linear solenoid and the actual valve timing change rate VTa, wherein a solid line curve shows the characteristic curve when the discharge pressure of working oil is relatively high while a broken line curve represents the characteristic curve when the discharge pressure of working oil is relatively low. In this conjunction, it is to be mentioned that the discharge pressure of working oil can vary in dependence on the engine rotation number NE and the temperature such as the cooling water temperature W.
As can be seen in FIG. 26, the reference current value ib changes constantly in dependence on the change in the discharge pressure of working oil. For instance, when the discharge pressure of working oil becomes low, the reference current value ib increases. Additionally, the manner in which the reference current value ib and hence the characteristic curve vary differs from one to another product such as the spool valve element 82 due to unevenness or dispersion of the dimensional factors or the like. When the discharge pressure of working oil becomes low, the rate of change VTa in the actual valve timing relative to the change of the linear solenoid current i will decrease.
Hereinafter, the linear solenoid current ib at which the actual valve timing Ta can remain invariable will be referred to as the holding current ih.
Ordinarily, when the valve timing is to be advanced with reference to the holding current ih, the linear solenoid current i may be set at a large value. On the contrary, when the valve timing is to be retarded, the linear solenoid current i may be set at a small value.
Next, valve timing detecting operation will be described by reference to FIG. 27 which is a timing chart for illustrating the crank angle signal SGT, the cam angle signal SGCd in the most retarded phase and the cam angle signal SGCa in the advanced phase. Phase relations between the crank angle signal SGT and the cam angle signals SGCd and SGCa as well as the actual valve timing Ta can be arithmetically determined on the basis of the data derived from the timing chart illustrated in FIG. 27.
The electronic control unit 100 is so designed or programmed as to measure a period T of the crank angle signal SGT as well as a phase-difference-equivalent time (hereinafter referred to as the phase difference time) .DELTA.Ta intervening between the cam angle signal SGCa and the crank angle signal SGT, i.e., the time corresponding to difference in phase between the cam angle signal SGCa and the crank angle signal SGT.
Further, the most retarded valve timing Td is arithmetically determined on the basis of a phase difference time ATd and the period T of the crank angle signal SGT in accordance with the following expression (1) when the valve timing retard is at maximum. EQU Td=(.DELTA.Td/T).times.180 [.degree. CA] (1)
The result of the calculation mentioned above is stored in a random access memory or RAM incorporated in the electronic control unit 100.
Further, the electronic control unit 100 is programmed or designed to determine arithmetically an actual valve timing Ta on the basis of a phase difference time .DELTA.Ta, the period T of the crank angle signal SGT and the most retarded valve timing Td in accordance with the following expression (2): EQU Ta=(.DELTA.Ta/T).times.180[.degree. CA]-Td (2)
Furthermore, the electronic control unit 100 is so designed as to make the actual valve timing Ta converge to a desired (or target) valve timing To through a feedback control of the linear solenoid current i on the basis of timing deviation or difference ER between the actual valve timing Ta and the desired valve timing To.
FIG. 28 is a block diagram showing schematically an internal configuration of the electronic control unit 100. As can be seen in the figure, the electronic control unit 100 includes a microcomputer 101.
Referring to FIG. 28, the microcomputer 101 is comprised of a CPU (central processing unit) 102 for executing various arithmetic operations, decision processings and others, a ROM (read-only memory) 103, a RAM (random access memory) 104 for storing temporarily the results of arithmetic operations (and/or other processings) executed by the CPU 102, an A/D (analogue-to-digital) converter 105 for converting an analogue signal to a digital signal, a counter 106 for counting the period of an input signal and/or other signal, a timer 107 for measuring a driving duration (duty cycle) of an output signal, an output port 108 constituting an output interface, and a common bus 109 for interconnecting the various blocks or components 102 to 108.
Provided in association with the microcomputer 101 is a first input circuit 110 which shapes the waveforms of the crank angle signal SGT outputted from the crank angle sensor 6 and the cam angle signal SGC generated by the cam angle sensor 24, wherein the output signal of the first input circuit 110 is supplied to the microcomputer 101 as an interrupt command signal INT.
Every time the interruption occurs in response to the interrupt command signal INT, the CPU 102 reads the value of the counter 106 to store it in the RAM 104.
Further, the CPU 102 arithmetically determines the period T of the crank angle signal SGT (see FIG. 27) on the basis of difference between the counter value at the time point when the preceding crank angle signal SGT was inputted and the current counter value, to thereby determine the engine rotation number (rpm) NE on the basis of the period T of the crank angle signal SGT.
Furthermore, the CPU 102 reads out from the RAM 104 the counter value in response to the cam angle signal SGC as inputted, to thereby determine arithmetically a phase difference time .DELTA.T on the basis of the deviation from the counter value at the time point the crank angle signal SGT was inputted.
Further provided in association with the microcomputer 101 is a second input circuit 111 for fetching the cooling water temperature W from the water temperature sensor 12, the throttle opening degree .theta. from the throttle sensor 27 and the intake air flow Q from the intake air-flow sensor 28, respectively, wherein the output signal of the second input circuit 111 undergone noise elimination processing, amplification and other processings is supplied to the A/D converter 105 which converts the signals representing the cooling water temperature W, the throttle opening degree .theta. and the intake air flow Q into corresponding digital data, respectively. The digital output data signals outputted from the A/D converter 105 are inputted to the CPU 102.
The driving circuit 112 is designed to output a control signal for driving the fuel injector 30, while the driving circuit 113 is designed to output a control signal for driving the ignitor 11.
In response to the various input signals, the CPU 102 arithmetically determines the driving time or duration for the fuel injector 30 as well as the ignition timing for the ignitor 11 on the basis of the input signals, while driving the fuel injector 30 and the ignitor 11 by way of the output port 108, the driving circuits 112 and 113, respectively, for thereby controlling the fuel injection quantity and the ignition timing.
The current control circuit 114 is designed to control the linear solenoid current i of the oil control valve 80. To this end, the CPU 102 determines arithmetically the value of the linear solenoid current i of the oil control valve 80 on the basis of the various input signals mentioned above to thereby output through the output port 108 a duty signal corresponding to the linear solenoid current i for the oil control valve 80 on the basis of the result of the time measurement performed by the timer 107.
On the other hand, the current control circuit 114 controls the linear solenoid current i flowing through the linear solenoid 83 of the oil control valve 80 in accordance with the duty signal mentioned above, to thereby realize the control of the valve open/close timing.
Further provided is a power circuit 115 which is designed to generate a constant voltage from the voltage of a battery 116 supplied via a key switch 117. Thus, the microcomputer 101 can operate with the constant voltage supplied from the power circuit 115.
In the conventional valve timing control system for the internal combustion engine known heretofore such as disclosed in the Japanese Unexamined Patent Application Publication No. 299876/1994 (JP-A-6-299876), learning of the most retarded valve timing Td is carried out under predetermined condition, e.g. in the idling state in which the engine is operated at a low speed.
Consequently, once the most retarded valve timing Td is learned erroneously, the learned value can not be corrected to the optimal value unless the idling state is resumed. For example, in the case where foreign material is undesirably deposited on the contacting portion(s) of the valve in the course of idling operation of the engine or where the valve is stopped on the way of displacement without being controlled to the regular or correct value, erroneous learning of the most retarded valve timing Td can be taken place.
When such erroneously learned state occurs, it takes tremendously much time to relearn the most retarded valve timing Td and resume the normal or regular value therefor when the engine is operating in a cruising mode (i.e., off-idle operation mode). Thus, the valve timing control can not be performed correctly before the learned value resumes the regular or correct value, which may incur not only degradation of engine operation performance but also deterioration in quality of the exhaust gas.
As will now be appreciated from the foregoing description, the conventional valve timing control system for the internal combustion engine suffers a problem that the engine operation performance as well as the exhaust gas quality is degraded because the most retarded valve timing Td can be learned only in the predetermined engine operation mode (e.g. idle mode of low engine speed) and because lots of time is taken for effectuating again the proper value for the most retarded valve timing Td once it has been learned incorrectly.